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United States Patent |
5,706,666
|
Yamanaka
,   et al.
|
January 13, 1998
|
Refrigeration apparatus
Abstract
A centrifugal type separator is disposed on a downstream side of a
temperature-operating type expansion valve. Liquid-phase refrigerant
separated herein is again pressure-reduced by an aperture resistance and
thereafter inducted to an inlet side of the evaporator by means of a
liquid-phase refrigerant discharge passage. Meanwhile, gas-phase
refrigerant separated by the separator is returned directly from a
gas-phase refrigerant discharge passage to an evaporator outlet side
passage, and after being united with superheated gas-phase refrigerant
evaporated by the evaporator, is taken into a compressor. A
temperature-sensing tube of the expansion valve is disposed further on the
downstream side of the foregoing union location so as to enable the
temperature of the superheated gas-phase refrigerant evaporated by the
evaporator to be sensed accurately.
Inventors:
|
Yamanaka; Yasushi (Nakashima-gun, JP);
Kakehashi; Nobuharu (Anjo, JP);
Kishita; Hiroshi (Anjo, JP);
Fujiwara; Kenichi (Kariya, JP)
|
Assignee:
|
Nippondenso Co., Ltd. (Kariya, JP)
|
Appl. No.:
|
796861 |
Filed:
|
February 6, 1997 |
Foreign Application Priority Data
| Apr 12, 1994[JP] | 6-73325 |
| Sep 22, 1994[JP] | 6-228112 |
Current U.S. Class: |
62/225; 62/512; 236/92B |
Intern'l Class: |
F25B 041/04 |
Field of Search: |
62/225,512
236/92 B
|
References Cited
U.S. Patent Documents
3738573 | Jun., 1973 | Eschbaugh et al. | 236/92.
|
4341086 | Jul., 1982 | Ishii | 62/225.
|
4370868 | Feb., 1983 | Kim et al. | 62/504.
|
4972676 | Nov., 1990 | Sakai | 62/512.
|
5056329 | Oct., 1991 | Wilkinson | 62/512.
|
Foreign Patent Documents |
2650935 | May., 1978 | DE.
| |
5-18635 | Jan., 1993 | JP.
| |
5-79725 | Mar., 1993 | JP.
| |
Primary Examiner: Tapolcai; William E.
Attorney, Agent or Firm: Cushman, Darby & Cushman IP Group of Pillsbury Madison & Sutro LLP
Parent Case Text
This is a division of application Ser. No. 08/420,490, filed Apr. 12, 1995,
now U.S. Pat. No. 5,619,861.
Claims
What is claimed is:
1. A temperature-opening type expansion valve comprising;
a body case;
a liquid-phase refrigerant influx passage provided on said body case into
which flows liquid-phase refrigerant condensed by a condenser;
a restricting passage provided on a downstream side of said liquid-phase
refrigerant influx passage within said body case to reduce pressure of
liquid-phase refrigerant;
a valve provided within said body case to regulate a degree of opening of
said restricting passage;
a gas-liquid separation means provided within said body case to separate
gas-liquid two-phase refrigerant pressure-reduced by said restricting
passage into liquid-phase refrigerant and gas-phase refrigerant;
a liquid-phase discharge passage provided in said body case so as to be
communicated with an inlet side of an evaporator, to discharge
liquid-phase refrigerant separated by the said gas-liquid separation means
to an inlet side of the evaporator;
a gas-phase discharge passage provided in said body case to discharge
gas-phase refrigerant separated by the said gas-liquid separation means to
an evaporator outlet side;
a gas-phase refrigerant passage provided in said body case so as to cause
gas-phase refrigerant evaporated by said evaporator to flow in and be
united with gas-phase refrigerant from said gas-phase refrigerant
discharge passage during a process thereof, and to cause gas-phase
refrigerant after said union to flow into an intake side of said
compressor;
a temperature-sensing means disposed in a location in said gas-phase
refrigerant passage which is upstream of a union location of gas-phase
refrigerant evaporated by said evaporator and gas-phase refrigerant from
said gas-phase refrigerant discharge passage, to sense temperature of
gas-phase refrigerant evaporated by said evaporator; and
a valve operating means to regulate a degree of opening of said valve in
correspondence to said gas-phase refrigerant temperature sensed by means
of said temperature-sensing means.
2. An expansion valve according to claim 1, wherein a passage is defined in
said body case such that gas-phase refrigerant from the evaporator may
flow thereinto, said passage communicating with said gas-phase discharge
passage.
3. Art expansion valve according to claim 1, wherein said gas-liquid
separation means comprises a centrifugal type separator.
4. Art expansion valve according to claim 3, wherein said gas phase
refrigerant passage is defined in said separator.
5. An expansion valve according to claim 4, wherein said gas phase
refrigerant passage is defined by an introduction pipe.
6. An expansion valve according to claim 1, wherein said body case is
constructed and arranged to be coupled to a condenser and an evaporator.
Description
CROSS REFERENCE TO RELATED APPLICATIONS
This application is based upon and claims priority from Japanese Patent
Application No. Hei 6-73325 filed Apr. 12, 1994 and Japanese Patent
Application No. Hei 6-228112 filed Sep. 22, 1994, with the contents of
each document being incorporated herein by reference.
BACKGROUND OF THE INVENTION
1. Field of the Invention
The present invention relates generally to a modification of a
refrigeration apparatus, for example suitable for employment in an
air-conditioning device for automotive use, and more particularly, to a
modification of an evaporator in a refrigeration apparatus for the purpose
of performance enhancement.
2. Description of the Related Art
An example provided in Japanese Patent Application Laid-open No. Hei
5-18635 exists as a prior example intended to enhance evaporator
performance in a refrigeration apparatus of this type. This prior art
disposes a gas-liquid separation chamber to separate gas-phase and
liquid-phase refrigerant on a core portion side end portion of a laminate
type evaporator composed by laminating flat tubing and corrugated fins,
providing in this gas-liquid separation chamber an inlet chamber to which
a refrigerant inlet pipe is connected and an outlet chamber to which a
refrigerant outlet pipe is connected, disposing an inlet side tank portion
communicated with an inlet tank of the core portion on a bottom portion of
the inlet chamber, disposing an outlet side tank portion communicated with
an outlet tank of the core portion on a bottom portion of the outlet
chamber, and further structured to cause the inlet chamber and the outlet
chamber to be communicated by a bypass passage portion at an uppermost
portion thereof.
By means of this, refrigerant of which the pressure has been reduced by a
pressure-reducing means of an expansion valve or the like to assume a
gas-liquid two-phase state is separated in a vertical direction by means
of specific gravity differential in the gas-liquid separation chamber;
liquid-phase refrigerant with high specific gravity is caused to flow from
the inlet chamber bottom portion through the inlet side tank portion and
into an inlet tank of the core portion, such that the liquid-phase
refrigerant is distributed uniformly of this inlet tank to multiple flat
tubing. Meanwhile, in the inlet chamber of the gas-liquid separation
chamber, gas-phase refrigerant with low specific gravity shifts to the
upper side, and passes through the bypass passage portion of the uppermost
portion to flow directly into (bypass) the outlet chamber. Accordingly,
gas-phase refrigerant which has exchanged heat with blown air for
air-conditioning use or the like and evaporated in the flat tubing passes
through the core portion outlet tank and flows into the outlet chamber.
Consequently, gas-phase refrigerant which has evaporated in the core
portion and gas-phase refrigerant which has bypassed from the gas-liquid
separation chamber is mixed in this outlet chamber, is discharged
externally from the outlet pipe, and is taken into the compressor.
As an incidental comment, according to the foregoing conventional device,
if gas-phase refrigerant can be separated sufficiently in the gas-liquid
separation chamber, this separated gas-phase refrigerant can be
distributed uniformly to the respective tubing of the core portion, and so
effective utilization for the purpose of heat exchange with the entirety
of the core portion without generating excessive or insufficient
refrigerant among the multiple tubing is possible.
However, specific experimentation and investigation by the inventors with
regard to the foregoing conventional device revealed the occurrence of the
problems that will be described hereinafter.
To wit, firstly, because gas-phase refrigerant which has been separated in
the gas-liquid separation chamber utilizing the specific gravity
differential of gas-phase and liquid-phase refrigerant is caused to flow
into the core portion inlet tank without change, at times such as during a
summer period when the cooling load is large, in a case of R134a
refrigerant, high pressure (i.e., pressure of the high-pressure circuit
from the compressor discharge side to the pressure-reducing means inlet
side of the refrigeration apparatus) assumes a high pressure of about 15
kg/cm.sup.2, and as a result thereof refrigerant downstream of the
pressure-reducing means (evaporator inlet) is subsequent to pressure
reduction, and so the degree of dryness thereof becomes large, a large
quantity of gas is generated, and the proportion of gas-phase refrigerant
in terms of the weight ratio reaches 40%. For this reason, gas-phase and
liquid-phase refrigerant cannot be separated sufficiently in the
gas-liquid separation chamber, and refrigerant flows into the core portion
in a state wherein gas-phase refrigerant is intermixed with the
liquid-phase refrigerant. It was discovered that, by means of this, the
distribution of liquid-phase refrigerant to the respective tubing in the
core portion becomes nonuniform, leading to a drop in evaporator
performance.
Secondly, in a case where a temperature-operating type expansion valve is
used as a pressure-reducing means, a temperature-sensing tube is disposed
on a downstream side of the refrigerant outlet pipe of the evaporator, and
so the temperature of refrigerant in which is mixed superheated gas-phase
refrigerant evaporated in the core portion and saturated gas-phase
refrigerant bypassed from the foregoing gas-liquid separation chamber is
necessarily detected. As a result of this, a temperature lower than the
actual temperature of the superheated gas-phase refrigerant of the
evaporator outlet by an amount corresponding to the saturated gas-phase
refrigerant comes to be detected, and the problem occurs wherein the
expansion valve cannot optimally regulate refrigerant flow to the
evaporator. Actually, it was discovered that the expansion valve tends to
close, leading to the problem of a drop in evaporator capacity.
Furthermore, when cooling load is small and the degree of dryness of
refrigerant downstream of the expansion valve is small, liquid-phase
refrigerant comes to be intermixed with gas-phase refrigerant separated in
the gas-liquid separation chamber and bypassed, and so the detected
temperature of the temperature-sensing tube drops further, and as a result
thereof the above-described problem becomes more marked.
SUMMARY OF THE INVENTION
In view of the above-described points, it is an object of the present
invention to provide a refrigeration apparatus which can enhance
uniformity of refrigerant distribution to respective tubing of an
evaporator even in a state of large cooling load.
Additionally, it is another object of the present invention to provide a
refrigeration apparatus disposing a refrigerant gas-liquid separation
means on a downstream side of a temperature-operating type expansion
valve, wherein a temperature-sensing means (temperature-sensing tube or
the like) of an expansion valve can accurately sense temperature of
superheated gas-phase refrigerant of an evaporator outlet, and which can
optimally control refrigerant flow to the evaporator.
In order to attain the foregoing objects, the present invention employs
technical means that will be described hereinafter.
A preferred mode of refrigerant apparatus in the invention includes a
compressor to compress and discharge refrigerant, a condenser to cool and
condense high-temperature, high-pressure gas-phase refrigerant discharged
from this compressor, a pressure-reducing means to reduce pressure of
liquid-phase refrigerant condensed by this condenser, a gas-liquid
separation means to separate gas-liquid two-phase refrigerant
pressure-reduced by this pressure-reducing means into liquid-phase
refrigerant and gas-phase refrigerant, a liquid-phase refrigerant
discharge passage to discharge liquid-phase refrigerant separated by this
gas-liquid separation means from the gas-liquid separation means, a
gas-phase refrigerant discharge passage to discharge gas-phase refrigerant
separated by this gas-liquid separation means from the gas-liquid
separation means, an auxiliary pressure-reducing means provided in this
liquid-phase refrigerant discharge passage to again reduce pressure of
refrigerant, an evaporator connected to a downstream side passage of this
auxiliary pressure-reducing means so that refrigerant again
pressure-reduced by this auxiliary pressure-reducing means flows in, to
evaporate this inflowing refrigerant, and an evaporator outlet side
passage to cause gas-phase refrigerant from this gas-phase refrigerant
discharge passage to unite with gas-phase refrigerant evaporated by this
evaporator and to be caused to be taken in on an intake side of the
compressor.
One of other preferred mode of the refrigerant apparatus includes a
compressor to compress and discharge refrigerant, a condenser to cool and
condense high-temperature, high-pressure gas-phase refrigerant discharged
from this compressor, a temperature-operating type expansion valve having
a restricting passage to reduce pressure of liquid-phase refrigerant
condensed by this condenser and a valve to regulate a degree of opening of
this restricting passage, a gas-liquid separation means to separate
gas-liquid two-phase refrigerant pressure-reduced by this
temperature-operating type expansion valve into liquid-phase refrigerant
and gas-phase refrigerant, a liquid-phase refrigerant discharge passage to
discharge liquid-phase refrigerant separated by this gas-liquid separation
means from the gas-liquid separation means, a gas-phase refrigerant
discharge passage to cause to discharge gas-phase refrigerant separated by
the gas-liquid separation means from the gas-liquid separation means, an
evaporator connected to a downstream side passage of this liquid-phase
refrigerant discharge passage so that refrigerant flows in from the
liquid-phase refrigerant discharge passage, to evaporate this inflowing
refrigerant, and an evaporator outlet side passage to cause gas-phase
refrigerant from the gas-phase refrigerant discharge passage to unite with
gas-phase refrigerant evaporated by this evaporator and to be caused to be
taken in on an intake side of the compressor wherein the
temperature-operating type expansion valve further comprises
temperature-sensing means disposed in locations in the evaporator outlet
side passage which are upstream of a union location of gas-phase
refrigerant evaporated by the evaporator and gas-phase refrigerant from
the gas-phase refrigerant discharge passage, to sense temperature of
gas-phase refrigerant evaporated by the evaporator, and
needle valve operating means to regulate a degree of opening of the needle
valve in correspondence to the gas-phase refrigerant temperature sensed by
these temperature-sensing means.
In an invention according to the preferred mode, after gas-liquid two-phase
refrigerant pressure-reduced by the pressure-reducing means is separated
into gas and liquid by the gas-liquid separation means, the refrigerant is
again pressure-reduced by the auxiliary pressure-reducing means provided
in the liquid-phase refrigerant discharge passage, and so the proportion
of refrigerant flow discharged from the gas-liquid separation means to the
liquid-phase refrigerant discharge passage side and of the refrigerant
flow discharged from the gas-liquid separation means to the gas-phase
refrigerant discharge passage side connected established so as to become
exactly the proportion of gas and liquid by means of the auxiliary
pressure-reducing means. As a result of this, even during high load
wherein high pressure becomes high, gas-phase refrigerant flowing into the
evaporator is made insignificant, uniformity of refrigerant distributed to
the respective tubing in the evaporator can be improved, and evaporator
performance can be effectively improved.
In an invention according to other preferred mode, when providing the
gas-liquid separation means on the downstream side of the
temperature-operating type expansion valve, the installation location of
the temperature-sensing means of the foregoing temperature-operating type
expansion valve is established at a location in the foregoing evaporator
discharge side passage which is upstream of the union location of
gas-phase refrigerant evaporated by the foregoing evaporator and gas-phase
refrigerant from the foregoing gas-phase refrigerant discharge passage,
and so superheated gas-phase refrigerant temperature from the evaporator
can be sensed accurately by means of the temperature-sensing means without
the saturated gas-phase refrigerant temperature from the gas-phase
refrigerant discharge passage being affected, and consequently the
refrigerant flow control effect of the expansion valve can be favorably
maintained.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic refrigeration apparatus cycle view indicating a first
embodiment according to the present invention;
FIGS. 2A , 2B and 2C are Mollier diagrams of respectively differing load
conditions;
FIG. 3 is a graph indicating a relationship between high pressure and
refrigerant dryness of a refrigeration apparatus;
FIG. 4 is a partially vertical sectional view of an expansion valve
indicating a second embodiment according to the present invention;
FIG. 5 is a perspective view of the expansion valve of FIG. 4;
FIG. 6A is a vertical sectional view of an expansion valve section
indicating a sectional view taken along line I--I of front view FIG. 6B;
FIG. 6B is a front view of the expansion valve;
FIG. 6C is a sectional view taken along line II--II of FIG. 6B;
FIG. 6D is a perspective view of the expansion valve section;
FIG. 7 is a schematic view of a refrigeration cycle indicating a fourth
embodiment according to the present invention;
FIG. 8A is a graph indicating a relationship between high pressure and
refrigerant dryness at the evaporator inlet side in the fourth embodiment;
FIG. 8B shows cycle behavior of a refrigeration apparatus at A, B and C
points in the graph of FIG. 8A;
FIG. 9A is an operation characteristic diagram of the fourth embodiment;
FIG. 9B is an operation characteristic diagram of an ON-OFF control cycle;
FIGS. 10A and 10B are sectional views of an EPR employed in the fourth
embodiment;
FIG. 11 is a schematic refrigeration apparatus cycle view indicating a
fifth embodiment according to the present invention;
FIG. 12 is a schematic cycle view of the essentials of a refrigeration
apparatus indicating a sixth embodiment according to the present
invention;
FIG. 13 is an operation characteristic diagram of the sixth embodiment;
FIG. 14 is a sectional view of a variable-capacity compressor employed in a
seventh embodiment; and
FIG. 15 is a sectional view of a variable-capacity compressor employed in
the seventh embodiment during low capacity.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
Embodiments according to the present invention will be described
hereinafter with reference to the drawings.
(First Embodiment)
FIG. 1 is a schematic cycle view of a refrigeration apparatus are
automotive air-conditioning use. Numeral 1 is a compressor, and is driven
by means of an automotive engine (drive source) 2 via an electromagnetic
clutch (operation interruption means) 1a. Numeral 3 is a condenser which
cools and condenses high-temperature and high-pressure gas-phase
refrigerant discharged from the compressor 1 by cooling air (cooling
medium) blown by a fan not illustrated.
Numeral 4 is a receiver which accumulates liquid-phase refrigerant
condensed by the condenser 3 and inducts only liquid-phase refrigerant to
the outlet side thereof. Numeral 5 is a temperature-operating type
expansion valve constituting a pressure-reducing means to cause pressure
reduction and expansion of refrigerant from the receiver 4, and numeral 5a
is a temperature-sensing tube constituting a temperature-sensing means
thereof. Numeral 6 is an evaporator for automotive air-conditioning use
for the purpose of cooling and dehumidifying conditioned air blown within
a passenger compartment.
This evaporator 6 may have a known structure, being a structure of aluminum
or similar metal having good thermal conductivity soldered integrally,
having an inlet tank (refrigerant distribution means) 6a, outlet tank
(refrigerant collection means) 6b, multiple tubing of flat configuration
(heat exchange portion refrigerant through-flow means) 6c structured by
joining two metal plates, and corrugated fins (heat exchange surface area
expansion means) 6d joined among this tubing 6c.
Numeral 7 is a centrifugal type separator constituting a gas-liquid
separation means to separate gas-liquid two-phase refrigerant
pressure-reduced by the foregoing expansion valve 5 into gas and liquid,
and is disposed on the downstream side of the expansion valve 5
independently of the expansion valve 5. This centrifugal type separator 7
has an internal space of tubular configuration 7a the axis of which is
disposed vertically, and creates swirl flow in refrigerant flow by
offsetting influx refrigerant from the expansion valve from a center of
tubular configuration and causing it to flow tangentially with respect to
this space 7a, and separates gas-liquid phase refrigerant by means of
centrifugal force generated by this swirl flow. That is to say, by
centrifugal force due to swirl flow at the internal space of tubular
configuration 7a of the separator 7, liquid-phase with high specific
gravity is shifted to the outer peripheral side, and gas-phase refrigerant
with low specific gravity is shifted to the center portion, and
accordingly liquid-phase refrigerant is collected on the lower side and
gas-liquid phase refrigerant is separated.
Numeral 8 is a liquid-phase refrigerant discharge passage which causes to
discharge liquid-phase refrigerant separated by the separator 7 and
collected at the lower portion of the space, and is removed from the
bottom portion of the separator 7 and communicated with the inlet tank 6a
of the foregoing evaporator 6. Numeral 9 is a gas-phase refrigerant
discharge passage which causes to discharge gas-phase refrigerant
separated by the separator 7, and is removed from the center portion of
the upper portion of the separator 7. 10 is an evaporator outlet side
passage (or, in other words, a compressor intake passage), and causes
gas-phase refrigerant from the outlet tank 6b of the evaporator 6 and
gas-phase refrigerant from the gas-phase refrigerant discharge passage 9
to be united and taken into the compressor 1.
The temperature-sensing tube 5a of the foregoing expansion valve 5 is
disposed further on the upstream side of the union location of the
gas-phase refrigerant discharge passage 9 to the passage 10, so as to
detect the temperature of only superheated gas-phase refrigerant
evaporated by the evaporator 6. Herein, refrigerant of identical type as
refrigeration apparatus circulating refrigerant is enclosed within the
temperature-sensing tube 5a, the pressure of this enclosed refrigerant
changes in accordance with the temperature of the foregoing superheated
gas-phase refrigerant and acts on a diaphragm 5b of the expansion valve 5.
The degree of opening of a needle valve (not illustrated) of the expansion
valve 5 is changed by means of displacement of this diaphragm 5b, and
consequently according to the present embodiment the needle valve
operating means is structured by the diaphragm 5b.
Numeral 11 is an aperture resistance disposed in the liquid-phase
refrigerant discharge passage 8 and constituting an auxiliary
pressure-reducing means to again reduce pressure of refrigerant supplied
to the evaporator, and concretely it is structured by a orifice (the
sectional configuration of a restricting passage thereof being of linear
configuration), nozzle (the inlet portion sectional configuration of a
restricting passage thereof being of smooth arc configuration), or the
like with an inner diameter of approximately 3 mm. Numeral 12 is an
aperture resistance provided in the gas-phase refrigerant discharge
passage 9 for the purpose of regulating gas-phase refrigerant flow of this
passage 9, and concretely it is structured by an orifice, nozzle,
short-length capillary tube, or the like.
Next, a mode of operation according to the present embodiment with the
above-described structure will be described with reference to the Mollier
diagrams of FIGS. 2A, 2B, and 2C. FIG. 2A indicates a state of refrigerant
(R134a) of respective portions of the refrigeration apparatus during
intermediate load in a spring/autumn period. The high-temperature,
high-pressure gas-phase refrigerant of the compressor 1 outlet is point A,
and liquid-phase refrigerant condensed by the condenser 3 and accumulated
by the receiver 4 flows into the expansion valve 5. The refrigerant of the
expansion valve inlet is point B, and gas-liquid two-phase refrigerant
after pressure reduction by means of the expansion valve 5 is point C.
High pressure P1 during intermediate load in a spring/autumn period is
normally approximately 7 kg/cm.sup.2 (=0.80 MPa), and when there is a drop
from this high pressure P1 to low pressure P2 (2.5 kg/cm.sup.2 =0.35 MPa)
at the expansion valve 5, the refrigerant dryness x becomes 0.20, and gas
with a specific gravity of approximately 20% is generated, but gas-liquid
two-phase refrigerant after this pressure reduction is centrifugally
separated in the separator 7. This separated liquid-phase refrigerant is
point D, and passes from here through the liquid-phase refrigerant
discharge passage 8 to be again pressure-reduced by means of the aperture
resistance 11, reaching point E of low pressure P3 (2.0 kg/cm.sup.2 =0.3
MPa). Separated gas-phase refrigerant is point F of saturated gas-phase
refrigerant, and passes from here through the gas-phase refrigerant
discharge passage 9 to be pressure-reduced by means of the aperture
resistance 12, reaching point G. The dryness x of the refrigerant at the
foregoing point E is approximately 0.05, and influx refrigerant to the
evaporator 6 is substantially all liquid-phase refrigerant, and so the
entirety of the heat exchanger portion can be utilized effectively for the
purpose of cooling operation, and evaporator performance can be improved,
without generating nonuniformity in refrigerant distribution to the
respective tubing 6c in the evaporator 6.
In the above-described mode of operation, pressure within the separator 7
means of disposing the aperture resistance 11 in the liquid-phase
refrigerant discharge passage 8, and the flow proportions to the
respective passages 8 and 9 are regulated so that the gas-liquid separated
gas-phase refrigerant and liquid-phase refrigerant respectively flow to
the respective passages 8 and 9 in exactly the corresponding amounts.
Furthermore, the aperture resistance 12 of the gas-phase refrigerant
discharge passage 9 plays a role in preventing an excessive increase in
refrigerant flow bypassed through this passage 9, and is not necessarily
required.
Superheated gas-phase refrigerant evaporated by the evaporator 6 reaches
point H, and thereafter unites with saturated gas-phase refrigerant from
the foregoing passage 9 in the evaporator outlet side passage 10, and
after the degree of superheating drops somewhat and there is a shift to
point I, it is taken into the compressor 1.
Next, FIG. 2B indicates the refrigerant state of respective portions of the
refrigeration apparatus during high load in a summer period. Because high
pressure P4 normally rises to 15 kg/cm.sup.2 =approximately 0.25 MPa,
refrigerant of point C (low pressure P5=4 kg/cm.sup.2 =approximately 0.50
MPa) after pressure reduction by the expansion valve 5 assumes a dryness x
of approximately 0.35, and the ratio of the amount of gas increases, and
so a portion of the gas-phase refrigerant attempts to be discharged to the
liquid-phase refrigerant discharge passage 8 side.
However, because the orifice or nozzle structuring the aperture resistance
11 has a property wherein the through-flow resistance of gas increases
sharply in comparison with liquid, when the gas-phase refrigerant attempts
to pass through the aperture resistance 11, the through-flow resistance of
the liquid-phase refrigerant discharge passage 8 side increases, pressure
P5 within the separator 7 increases, and the amount of gas escaping to the
gas-phase refrigerant discharge passage 9 side increases. As a result of
this, the amount of gas intermixed in the liquid-phase refrigerant
discharge passage 8 side does not increases appreciably, and even in a
high-load condition wherein high pressure P4=approximately 15 kg/cm.sup.2
(=0.25 MPa), the degree of refrigerant dryness x of the evaporator inlet
(point E) can be suppressed to a smaller value than approximately 0.15,
and by means of this refrigerant distribution to the respective tubing 6c
in the evaporator 6 can be maintained in a favorable state, and
performance of the evaporator 6 can be assured. Furthermore, low pressure
P6 of point E is approximately 2 kg/cm.sup.2 (=0.3 MPa).
Next, FIG. 2C indicates the refrigerant state of respective portions of the
refrigeration apparatus during low load in a winter period. Because high
pressure P7 normally declines to approximately 5 kg/cm.sup.2, refrigerant
of point C (low pressure P8=2.3 kg/cm.sup.2 =approximately 0.33 MPa) after
pressure reduction by means of the expansion valve 5 assumes a degree of
dryness x of 0.1, and the ratio of the amount of gas decreases, and so a
portion of the liquid-phase refrigerant is discharged to the gas-phase
refrigerant discharge passage 9 side.
Meanwhile, because the degree of refrigerant dryness x of the refrigerant
at the evaporator inlet (point E) becomes a small value of approximately
0.02, refrigerant distribution to the evaporator tubing 6c naturally
becomes favorable, and evaporator performance can be assured.
Additionally, with respect to the compressor 1, refrigerant of moist vapor
of point G and superheated gas-phase refrigerant of point H are united,
and refrigerant of moist vapor of point I (refrigerant including a portion
of liquid) attempts to be taken into the compressor 1, and so return
performance of lubrication oil which is frequently a problem in winter
periods is improved, and reliability of the compressor 1 is improved.
Moreover, low pressure P9 of point E is approximately 2 kg/cm.sup.2.
FIG. 3 shows the relation between the dryness x of the refrigerant as the
vertical axis and high pressure P of the refrigeration apparatus as the
horizontal axis; solid line (1) indicates the dryness x of the refrigerant
of the evaporator inlet according to the present invention (point E of
FIG. 2). In FIG. 3, (a) indicates the dryness during intermediate load,
(b) indicates the dryness during high load, and (c) indicates the dryness
during low load. It is understood that, according to the present
invention, the dryness x can be constantly maintained at a small value,
regardless of load fluctuation.
In contrast to this, the dryness x of an evaporator inlet in an ordinary
refrigeration apparatus according to the prior art is, as shown by dotted
line (2), a much larger value compared with the apparatus according to the
present invention, and is a cause of evaporator performance deterioration.
Moreover, dotted dash line (3) indicates a dryness of refrigerant (point F
of FIG. 2) at the inlet of the gas-phase refrigerant discharge passage 9
in an apparatus according to the present invention; it is understood that
it is at a state proximate to saturated gas of a constant degree of
dryness x=1, regardless of load fluctuation.
Furthermore, as is understood from the foregoing description of the mode of
operation, the liquid-phase refrigerant discharge passage 8 and gas-phase
refrigerant discharge passage 9 in the present invention are not to be
interpreted as exclusively causing to discharge always only liquid-phase
refrigerant or gas-phase refrigerant, but refer to a case wherein a
portion of gas-phase refrigerant or liquid-phase refrigerant may be
intermixed due to fluctuation or the like in load conditions of the
refrigeration apparatus, compressor revolving speed, or the like.
(Second Embodiment)
FIGS. 4 and 5 indicate a second embodiment, structured by integrating a
centrifugal type separator 7 with the above-described expansion valve 5.
In FIGS. 4 and 5, numeral 13 is a body case of the expansion valve 5
formed in a substantially cubic configuration by metal of aluminum or the
like. The vertical direction of FIG. 2 corresponds to the vertical
direction if a state of actual use, and a refrigerant inlet 14 into which
liquid-phase refrigerant flows from a receiver 4 is opened on a lower
portion right-hand side of this body case 13.
This refrigerant inlet 14 is communicated with a needle valve housing
chamber 15 formed on a lower central portion of the body case 13, and a
needle valve 16 and valve spring 17 are housed within this chamber 15. The
installation load of this spring 17 is adjustable by means of an
installation plate 18 fixed by screw to the body case 13. Herein, a
liquid-phase refrigerant discharge passage of the expansion valve 5 is
structured by the refrigerant inlet 14 and needle valve housing chamber
15.
Numeral 19 is a restricting passage formed on the downstream side of this
liquid-phase refrigerant discharge passage for the purpose of reducing
pressure of liquid-phase refrigerant. The degree of opening of this
restricting passage 19 is adjusted by means of the needle valve 16.
According to the present embodiment, a centrifugal type separator 7 is
disposed immediately after this restricting passage 19, with this
separator 7 disposed in a substantially central portion in the vertical
direction of the body case 13. Accordingly, the direction of flow of
gas-liquid two-phase refrigerant after pressure reduction squirted out
from the restricting passage 19 is offset from the tubular center with
respect to an internal space of tubular configuration 7a of the separator
7 (see FIG. 5), and so this gas-liquid two-phase refrigerant generates a
swirl flow R in this internal space of tubular configuration 7a.
Numeral 20 is a gas-phase induction pipe disposed so as to protrude from a
central location of the internal space of tubular configuration 7a of the
separator 7, and is connected to a gas-phase refrigerant discharge passage
9. According to the present embodiment, this passage 9 itself is caused to
double in duty with the role of the aperture resistance 12 of FIG. 1 by
establishing the inner diameter of this passage 9 appropriately. An
evaporator outlet side passage 10 is formed in an upper portion of this
body case 13 so as to pass therethrough laterally in tubular
configuration, and an outlet portion of the gas-phase refrigerant
discharge passage 9 is united at a rightward (i.e., toward the outlet)
section in the lateral direction of this passage 10.
A temperature-sensing member 5c of the expansion valve 5 constitutes a
temperature-sensing means to sense temperature of superheated gas-phase
refrigerant evaporated by an evaporator 6, and is disposed at a location
in the evaporator outlet side passage 10 further on the upstream side than
the union location of the outlet portion of the foregoing gas-phase
refrigerant discharge passage 9 so that the temperature of the foregoing
superheated gas-phase refrigerant can be detected accurately. An outlet
end 10a of the passage 10 is connected to the intake side of a compressor
1, and an inlet end 10b is connected to an outlet tank 6b of the
evaporator 6.
Numeral 21 is a joint member constituting a passage connecting means, and
is formed in a configuration wherein first and second pipe portions 22 and
23 are formed integrally with an interconnecting plate 24 of a metal being
aluminum or the like. The first pipe portion 22 forms the liquid-phase
refrigerant discharge passage 8 of FIG. 1, and an aperture resistance 11
composed of an orifice is formed intermediately therein. This first pipe
portion 22 connects an opening end 7b of the internal space of tubular
configuration 7a of the centrifugal type separator 7 and an inlet tank 6a
of the evaporator 6, and according to the present is modified as will be
described below for the purpose of guiding centrifugally separated
liquid-phase refrigerant favorably to the foregoing aperture resistance 11
side.
Briefly, the diameter of the opening end 7b in the centrifugal type
separator 7 is enlarged with respect to the internal space of tubular
configuration 7a, and along with this, the center of the opening end 7b is
lowered to lower than the center of the space 7a (see FIG. 5), and
accordingly the location of the aperture resistance 11 is established so
as to oppose a section lower than the center of this opening end 7b. By
means of this, liquid-phase refrigerant centrifugally separated, shifted
to an outer peripheral side, and collected in the internal space of
tubular configuration 7a is collected at a lower portion of the opening
end 7b by means of gravity, and thereafter can be caused to flow smoothly
into the passage of the aperture resistance 11.
The second pipe portion 23 of the joint member 21 connects the outlet tank
6b of the evaporator 6 and the inlet end 10b of the evaporator outlet side
passage 10.
Next, to describe the operating mechanism of the needle valve 16 of the
expansion valve 5, the needle valve 16 is interconnected integrally with
an operating rod 25, an upper end of this operating rod 25 contacts the
temperature-sensing member 5c, and according to the present embodiment
this temperature-sensing member 5c is structured from a cylindrical body
formed of aluminum or similar metal having good thermal conductivity.
Accordingly, the upper end of this temperature-sensing portion 5c contacts
a diaphragm 26 disposed on an outer-surface side of an uppermost portion
of the body case 13, and so the needle valve 16 is also displaced via the
temperature-sensing member of cylindrical configuration 5c and operating
rod 25 in correspondence with vertical displacement of this diaphragm 26.
A space 27 on the lower side of the diaphragm 26 is communicated with the
evaporator outlet side passage 10 via a communicating passage 28 on the
perimeter of the temperature-sensing member 5c, and so pressure within the
space 27 is identical to the pressure of the passage 10. Meanwhile, a
space 29 on an upper side of the diaphragm 26 is sealed by a cover 30, and
moreover refrigerant of identical type as circulating refrigerant of the
refrigeration apparatus is enclosed in the interior thereof. This enclosed
gas conducts, via the metal diaphragm 26, the superheated gas-phase
refrigerant temperature of the evaporator outlet sensed by the
temperature-sensing member 5c, and exhibits pressure changes in
correspondence to this superheated gas-phase refrigerant temperature. The
diaphragm 26 is abundant in elasticity and has good thermal conductivity,
and is preferably formed with a tough material, being composed for example
of stainless steel or a similar metal.
Because the operating mechanism of the needle valve 16 of the expansion
valve 5 is structured in the above-described manner, the needle valve 16
is displaced by a balance of pressure corresponding to superheated
gas-phase refrigerant temperature compressing the diaphragm 26 downwardly,
refrigerant pressure of the passage 10 compressing the diaphragm 26
upwardly, and installation load of the spring 17, by means of which the
degree of opening of the restricting passage 19 is controlled so as to
maintain the degree of superheating of the gas-phase refrigerant of the
evaporator outlet at a predetermined value. Consequently, according to the
present embodiment, a needle valve operating means is structured by means
of the temperature-sensing member 5c, diaphragm 26, and the like.
As described above, the second embodiment indicated in FIGS. 4 and 5 is
characterized by a centrifugal type separator 7 built into the expansion
valve 5 as an integral structure, and is identical to the first embodiment
in other respects, and so because the mode of operation as the
refrigeration apparatus indicated in FIGS. 2 and 3 is similar for the
second embodiment as well, description thereof will be omitted.
(Third Embodiment)
FIGS. 6A through 6D indicates a third embodiment which is a slight
modification of the structure of the second embodiment. According to the
present embodiment, a gas-phase refrigerant discharge passage 9 is formed
by hole-drilling upwardly from a bottom surface of a body case 13, and so
an opening end below this passage 9 is sealed by a closing member 31.
(Fourth Embodiment)
FIGS. 7 to 10 describe a fourth embodiment applying the present invention
in a refrigeration apparatus for automotive air-conditioning use provided
with an evaporation pressure regulating valve (hereinafter termed an
"EPR").
According to the refrigeration apparatus for automotive air-conditioning
indicated in FIG. 1, in order to prevent a decline in cooling performance
due to frosting of an evaporator 6, blown air temperature of the
evaporator 6 is normally detected by a temperature sensor (not
illustrated) being a thermistor or the like, and ON-OFF control of running
of a compressor 1 is performed in accordance with the detected temperature
thereof.
That is to say, if blown air temperature drops to an established
temperature of for example 3.degree. C., electrical conduction of an
electromagnetic clutch 1a is interrupted and the compressor 1 is stopped,
thereby causing the temperature of the evaporator 6 to drop to 0.degree.
C. or less and preventing the occurrence of frosting. Accordingly, if
blown air temperature rises to an established temperature of for example
4.degree. C., ON-OFF control is performed to electrify the electromagnetic
clutch 1a and start the compressor 1.
According to experimentation and investigation by the inventor, it was
discovered that the problem which will be described next occurs in a case
of ON-OFF control of running of the foregoing compressor 1 according to
the structure of the first embodiment indicated in FIG. 1.
Briefly, FIG. 8A takes high-pressure pressure P of the cycle as the
horizontal axis and the refrigerant degree of dryness X of the evaporator
inlet as the vertical axis. In the drawing, A indicates a refrigerant
dryness X of the evaporator inlet under high-load conditions of evaporator
intake air (conditioned air) temperature: 35.degree. C., humidity: 60%,
amount of air: 480 m.sup.3 /h, and cycle low-pressure pressure: 0.3 MPa.
B indicates a refrigerant dryness X of the evaporator inlet under
intermediate-load conditions of evaporator intake air (conditioned air)
temperature: 27.degree. C., humidity: 50%, amount of air: 480 m.sup.3 /h,
and cycle low-pressure pressure: 0.3 MPa.
C indicates a refrigerant degree of dryness X of the evaporator inlet under
low-load conditions of evaporator intake air (conditioned air)
temperature: 25.degree. C., humidity: 30%, amount of air: 300 m.sup.3 /h,
and cycle low-pressure pressure: 0.3 MPa.
According to FIG. 8B, to describe cycle behavior with respect to the
broad-ranging conditions indicated by the foregoing A, B, and C, under the
high-load conditions of point A, the dryness X of the refrigerant
immediately after the expansion valve 5 (inlet of the evaporator 6) is
large, and so the generated amount of gas-phase refrigerant increases. For
this reason, even if the main stream of gas-phase refrigerant is bypassed
to the gas-phase refrigerant discharge passage 9, a portion of the
gas-phase refrigerant is intermixed into the liquid-phase refrigerant of
the liquid-phase refrigerant discharge passage 8 (see model A below FIG.
8A).
Next, under the intermediate-load conditions of point B, the degree of
dryness X of the refrigerant immediately after the expansion valve 5 is
reduced, the generated amount of gas-phase refrigerant becomes equal to
the amount of gas-phase refrigerant bypassed to the gas-phase refrigerant
discharge passage 9, and the two assume a balanced state, and so
intermixing of gas-phase refrigerant into the liquid-phase refrigerant of
the liquid-phase refrigerant discharge passage 8 disappears (see model B
below FIG. 8B).
Finally, under the low-load conditions of point B, the degree of dryness X
of the refrigerant immediately after the expansion valve 5 is reduced and
the generated amount of gas-phase refrigerant is reduced further, and so a
portion of the liquid-phase refrigerant is also bypassed to the gas-phase
refrigerant discharge passage 9 side. In this case, naturally, only
liquid-phase refrigerant flows on the liquid-phase refrigerant discharge
passage 8 side, and gas-phase refrigerant is not intermixed (see model C
below FIG. 8C).
According to a system establishing refrigerant flow ratios to the two
refrigerant passages 8 and 9 by fixed aperture resistances 11 and 12 in
this manner, under low-load conditions wherein high- and low-pressure
differential pressure is small and the amount of gas generated downstream
of the expansion valve 5 is small, bypassing of liquid-phase refrigerant
to the gas-phase refrigerant discharge passage 9 side cannot be avoided.
However, as shown in FIG. 9B, in an ON-OFF control cycle, high-pressure and
low-pressure pressures of the cycle fluctuate greatly each time ON-OFF
control for the compressor 1 is performed, and accordingly a region Z
wherein high- and low-pressure differential pressure drops excessively to
a predetermined value or less is generated during the initial starting
period and immediately after running stoppage of the compressor 1.
In this region Z, the cycle conditions of point C of the foregoing FIG. 8
are effected, and liquid-phase refrigerant discharge (bypass) to the
gas-phase refrigerant discharge passage 9 occurs. ON-OFF control of the
compressor 1 is performed frequently and repeatedly, and so the amount of
liquid-phase refrigerant bypass increases, causing the problem of an
increase in compressor drive power.
In order to solve this problem, reducing the aperture of the aperture
resistance 12 of the gas-phase refrigerant discharge passage 9 may be
considered, but if the aperture of this aperture resistance 12 is reduced,
it invites a decrease in the bypass amount of gas-phase refrigerant to the
gas-phase refrigerant discharge passage 9 during constant running of the
compressor 1, and the amount of decrease of the bypass amount of this
refrigerant causes the phenomenon of an increase in the amount of
gas-phase refrigerant to the liquid-phase refrigerant discharge passage 8
without change. This conflicts with the aim of causing a decrease in the
refrigerant degree of dryness at evaporator 6 inlet and achieving
uniformity in refrigerant distribution, which is an object of the present
invention, and is not preferred.
Accordingly, as a means for the purpose of preventing frosting of the
evaporator 6, the fourth embodiment disposes an EPR 40 on the downstream
side of the evaporator 6 as shown in FIG. 7. By means of regulating
refrigerant evaporation pressure at the evaporator 6 to a predetermined
value or more by means of the EPR 40, ON-OFF control of the compressor 1
is rendered unnecessary.
Concretely, this EPR 40 is structured as shown in FIGS. 10A and 10B; an
inlet 41a and outlet 41b are formed on a needle valve body 41, and a
refrigerant path 41c communicating the inlet 41a and outlet 41b is formed
in the interior thereof. In this refrigerant path 41c, a needle valve 42
of spool configuration for path opening/closing is disposed slidably in
the lateral directions of the drawing.
This needle valve 42 is constantly compressed in the right-hand direction
of the drawing (i.e., inlet 41a side) by means of a coil spring 43 so as
to contact and close a valve seat portion 41d of the needle valve body 41.
The spring 43 is housed within an expandable bellows 44. An inert gas (for
example N.sub.2 gas) is enclosed within this bellows 44 at a predetermined
pressure (for example 1 kg/cm.sup.2), minimizing the influence of
secondary pressure of the refrigerant path 41c on the opening/closing
operation of the needle valve 42.
Additionally, two guide members 45 and 46 are disposed on an inner
peripheral side of the spring 43; the guide member 46 is mated slidably
with an inner peripheral surface of the guide member 45. In addition, a
plurality of open portions for the purpose of causing refrigerant to flow
when the valve is open are provided on a portion of tubular configuration
of the needle valve 42 so as to open radially.
When refrigeration load (cooling load) becomes small and the refrigerant
evaporation pressure of the evaporator 6 (refrigerant pressure on the
inlet 41a side) drops, the needle valve 42 is compressed to the right
hand-side (i.e., the inlet 41a side) by the force of the spring 43 as is
shown in FIG. 10 A, and contacts the valve seat portion 41d of the needle
valve body to close the valve. By means of this, evaporation pressure is
maintained at a predetermined value.
Conversely, when refrigeration load (cooling load) becomes large and the
refrigerant evaporation pressure of the evaporator 6 (refrigerant pressure
on the inlet 41a side) rises, the needle valve 42 resists the force of the
spring 43 and is compressed to the left-hand side (i.e., the outlet 41b)
side, and is released from the valve seat portion 41d of the needle valve
body 41. By means of this, refrigerant flows through the path through the
open portions 42a of the needle valve 42, the EPR 40 assumes a valve-open
state, and refrigeration performance of the specified period is
demonstrated.
In this foregoing manner, the evaporation pressure of refrigerant is
maintained at not less than a predetermined value (for example, in a case
of R134a refrigerant, the equivalent of 0.3 MPa and 0.degree. C.
refrigerant evaporation temperature), and occurrence of frost on the
evaporator 6 can be prevented.
In FIG. 7. The gas-phase refrigerant discharge passage 9 is connected in
the evaporator outlet side passage 10 at an intermediate location
discharge of the disposed location of the temperature-sensing tube 5a of
the expansion valve 5 and upstream of the EPR 40, and additionally an
outer-average tube 5d of the expansion valve 5 is connected to the
downstream side of the EPR 40, to induct refrigerant pressure of the EPR
40 downstream side to a diaphragm 5b).
Consequently, the expansion valve 5 controls the needle valve degree of
opening by means of the refrigerant temperature of the evaporator 6 outlet
detected by means of the temperature-sensing tube 5a, the refrigerant
pressure of the EPR 40 downstream side, and a priorly established spring
(not illustrated; see the spring 17 of FIGS. 4, 5, and 6) installation
load.
According to the fourth embodiment, refrigerant evaporation pressure of the
evaporator 6 is controlled at not less than a predetermined value by means
of the EPR 40, and so the need to perform ON-OFF control of the compressor
1 for the purpose of frost prevention of the evaporator 6 is eliminated.
For this reason, as shown in FIG. 9A the compressor 1 is maintained
without change in an ON (operating) state, and so high-pressure pressure
and low-pressure pressure of the cycle are also maintained substantially
uniformly if the refrigeration (cooling) cycle is uniform. Consequently,
liquid-phase refrigerant bypassing to the gas-phase refrigerant discharge
passage 9 caused by a decline of the high- and low-pressure differential
pressure to the predetermined value or less does not occur.
Moreover, according to the structure of the fourth embodiment, there is no
particular need to reduce the aperture resistance 12 of the gas-phase
refrigerant discharge passage, and so the problem of an increased amount
of gas-phase refrigerant discharge to the liquid-phase refrigerant
discharge passage 8 during constant running of the compressor 1 also does
not occur.
(Fifth Embodiment)
In a cycle according to the fourth embodiment indicated in FIG. 7, the
outer-average tube 5d of the expansion valve 5 is connected to the EPR 40
downstream side, and during low-load conditions when the EPR 40 restricts
the degree of valve opening thereof to control evaporation pressure, the
pressure drop of the EPR 40 outlet side is inducted to the diaphragm 5b of
the expansion valve 5 by means of the outer-average tube 5d. By means of
this, foregoing low load the degree of opening of the expansion valve 5 is
forcibly caused to increase, liquid-phase refrigerant containing
lubrication oil is returned to the compressor 1, and insufficient
lubrication in the compressor 1 is prevented.
However, in a cycle according to the present invention, during low load
wherein the EPR 40 is actuated, liquid-phase refrigerant containing
lubrication oil can be caused to bypass to the gas-phase refrigerant
discharge passage 9 by means of adjusting the ratio of the two aperture
resistances 11 and 12, and insufficient compressor lubrication during low
load can be solved by means of this liquid-phase refrigerant bypass.
A fifth embodiment addresses this point; as shown in FIG. 11, an
outer-average tube 5d of an expansion valve 5 is connected to an inlet
side of an EPR 40, and in other respects the fifth embodiment is identical
to the fourth embodiment.
According to the fourth embodiment, an increase in refrigerant discharge
due to an increased degree of opening of the expansion valve 5 and an
increase in compressor power consumption on the basis thereof occur during
low load, but according to the fifth embodiment, the expansion valve 5
adjusts refrigerant flow so that the degree of refrigerant superheating of
the evaporator outlet assumes a predetermined value even during low load
when the EPR 40 is actuated, and so compressor power consumption can be
reduced without generating excessive refrigerant flow, and so the
performance coefficient of the refrigeration cycle can be improved.
(Sixth Embodiment)
In a further modification of the fifth embodiment, as shown in FIG. 12, the
gas-phase refrigerant discharge passage 9 is connected to an outlet side
of the EPR 40, and refrigerant from this passage 9 is caused to be
bypassed directly to downstream of the EPR 40.
FIG. 13 indicates the refrigerant pressures of respective portions a
through f of FIG. 12. The a through f of the horizontal axis of FIG. 13
indicate the respective portions a through f of FIG. 12.
The solid lines of FIG. 13 indicate the refrigerant pressures of the
respective portions when fully open (during high load), and the broken
lines indicate the refrigerant pressures of the respective portions during
EPR actuation (during low load). During EPR actuation, the pressure of
point f of the EPR outlet side drops as shown by the broken line due to a
path-restricting action by means of the needle valve 42 of the EPR 40.
The flow ratio of refrigerant to the liquid-phase refrigerant discharge
passage 8 and the gas-phase refrigerant discharge passage 9 is mainly
determined by the aperture of the aperture resistances 11 and 12, and
secondly is determined by the before and after differential pressures of
the respective aperture resistances 11 and 12. The sixth embodiment
addresses the before and after differential pressures of these respective
aperture resistances 11 and 12, causing liquid-phase refrigerant to be
bypassed effectively to the gas-phase refrigerant discharge passage 9 only
during low load, when liquid-phase refrigerant return to the compressor 1
is required.
Specifically, to describe a mode of operation of the sixth embodiment,
refrigerant flow within the cycle increases during high load, and so
refrigerant return to the compressor 1 is unnecessary. During this high
load, the needle valve 42 of the EPR 40 is opened fully, and so the
pressure of the outlet point f of the EPR 40 is substantially identical to
the inlet point d of the evaporator 6.
Consequently, the before and after differential pressure .DELTA.PM of the
aperture resistance 11 of the liquid-phase refrigerant discharge passage 8
and the before and after differential pressure .DELTA.PB1 of the aperture
resistance 11 of the gas-phase refrigerant discharge passage 9 come to be
substantially identical. In a condition wherein these two before and after
differential pressures are substantially identical, the aperture ratio of
the aperture resistances 11 and 12 is established priorly so that
liquid-phase refrigerant is not bypassed to the gas-phase refrigerant
discharge passage 9 side, and thereby bypass of liquid-phase refrigerant
to the gas-phase refrigerant discharge passage 9 side does not occur.
Meanwhile, during low load, the pressure of point f of the EPR outlet side
drops as shown by the broken line due to a path-restricting action by
means of the a 42 of the EPR 40. Because of this, the before and after
differential pressure of the aperture resistance 12 of the gas-phase
refrigerant discharge passage 9 increases as indicated by .DELTA.PB2. For
this reason, bypass flow to the gas-phase refrigerant discharge passage 9
increases, and liquid-phase refrigerant can be caused to be bypassed to
the passage 9 side.
As a result of this, according to the sixth embodiment, the phenomenon
whereby pressure of point f of the EPR 40 outlet drops at the time of a
low-load condition wherein the EPR 40 is actuated is utilized, and
liquid-phase refrigerant return can be achieved only when liquid-phase
refrigerant return is required, without adding a special structure.
(Seventh Embodiment)
The above-mentioned fourth through sixth embodiments described a case
wherein an EPR 40 is utilized as a means of preventing frost formation on
the evaporator 6, but a seventh embodiment utilizes as the compressor 1 a
variable-capacity type which can continuously vary the discharge capacity
thereof, maintaining evaporation pressure at a predetermined value or more
by means of continuously controlling the capacity of the compressor 1 so
as to prevent frost formation of the evaporator 6.
That is to say, FIG. 9A indicates compressor 1 ON-OFF and high- and
low-pressure pressures in a case of an EPR cycle, but if the refrigeration
(cooling) load is stable with the compressor 1 remaining on, similarly to
FIG. 9A, high pressure and low pressure are maintained uniformly even in a
cycle employing a compressor 1 of variable-capacity type.
FIGS. 14 and 15 indicate an example of a compressor 1 of continuously
variable capacity type; numeral 101 is a rotating shaft which receives
drive force from an engine to rotate. This rotating shaft 101 is
instructed freely and rotatably to a housing via bearings 102 and 103.
An inclined plate 104 is installed on the rotating shaft 101 so that the
tilt angle thereof can be varied. That is to say, the rotational center
location of the inclined plate 104 is freely rotatable at a spherical
support portion 105, and moreover is mated with a two-faced width portion
107 within a groove 106 formed on the inclined plate 104 side, so that
rotation of the rotating shaft 101 is conveyed to the inclined plate 104.
Additionally, a pin 108 is fixed to the inclined plate 104 via the groove
portion 106, so that the tilt angle of the inclined plate 104 is varied by
movement of this pin 108 within a long groove 109 formed in the two-faced
width portion 107 of the rotating shaft 101.
The inclined plate 104 is interconnected with a piston via a shoe 110, and
the piston 111 receives rocking motion of the inclined plate 104 to slide
reciprocatingly within a cylinder 112. In an intake stroke in which the
volume of an operating chamber 113 expands in accompaniment to the
reciprocating sliding of this piston 111, an intake valve 114 opens and
refrigerant is taken into the operating chamber 113 side from an intake
chamber 115. Meanwhile, in a discharge stroke in which the volume of the
operating chamber 113 decreases in accompaniment to the reciprocating
sliding of this piston 111, refrigerant is discharge through a discharge
valve 116 to a discharge chamber 117. Moreover, the intake chamber 115 is
communicated with an intake port 118 via an intake passage within the
compressor 1 so that low-temperature, low-pressure refrigerant taken in
from the evaporator 6 of the refrigeration cycle is supplied. Meanwhile,
the discharge chamber 117 is communicated with a discharge port 119
through discharge passage within the compressor 1 so that refrigerant is
discharged to the condenser 3 side of the refrigeration cycle from the
discharge port 119 thereof.
The discharge volume of this compressor 1 is continuously varied by means
of variable control of the reciprocating stroke amount of the piston 111.
Variation of the reciprocating stroke amount of this piston 111 is
performed by means of varying the tilt angle of the inclined plate 104.
Variation of this tilt angle is performed by means of interlocking and
varying the rotational center location and tilt angle of the inclined
plate 104 in a state wherein this of constantly uniform top dead center
indicated in the center right-hand side of FIG. 14.
According to the present embodiment, the foregoing control is performed by
means of employing a spool 120 to cause the spherical support portion 105
to be displaced along the rotating shaft 101 in the lateral direction of
the center of the drawing. The locational displacement of the spool 120 is
performed by means of regulating pressure within a control-pressure
chamber 121 caused to be formed on a rear surface thereof. That is to say,
one side of the spool 120 becomes the intake chamber 115 to which intake
pressure is constantly applied. In contrast to this, the control chamber
121 is supplied with pressure which has been adjusted by means of a
control valve 122, and differential pressure of pressure within this
control-pressure chamber 121 and pressure within the intake chamber 115 is
applied to the spool 120. Accordingly, the location of the tilt angle is
controlled at a location balanced by pressure applied to this spool 120
and the compression reaction force with the piston 111.
Furthermore, the control valve 122 adjusts discharge pressure supplied from
the discharge chamber 117 through a high-pressure induction passage 123
and low pressure (intake pressure) supplied from a low-pressure induction
passage 124, supplying uniform pressure from a control-pressure passage
125 to the control-pressure chamber 121, and so according to the present
embodiment an electrical-control type which switches the foregoing two
passages 123 and 124 by means of electrical signals is utilized.
It is also possible to employ a pressure-responsive member such as a
diaphragm as the control valve 122 to utilize a structure to regulation
control pressure by means of a purely mechanical mechanism.
FIG. 14 indicates a state wherein a predetermined pressure has been
supplied to the control-pressure chamber 121 and the spool 120 has been
shifted to the left-hand side of the center of drawing by a predetermined
amount.
FIG. 15 indicates a state wherein the discharge capacity of the compressor
1 has been reduced further from the state indicated in this FIG. 14. In
this state, intake pressure is supplied to the control-pressure chamber
121. As a result of this, the spool 120 is displaced by the maximum amount
to the right-hand side of the center of the drawing in accompaniment with
compression reaction force and the like of the piston 111. As a result of
this, the rotational center location of the inclined plate 104 is also
displaced to the right-hand side of the center of the drawing, and the
tilt angle of the inclined plate 104 is also displaced in a direction
approaching a right angle with respect to rotating shaft 101.
As is clear from FIG. 15, in this state the amount of rocking of the
inclined plate 104 is small as well, and consequently the reciprocating
stroke of the piston 111 is at a minimum.
In a cycle employing a variable-capacity compressor 1, generally, during
low load (i.e., when high pressure is low), the capacity of the compressor
1 becomes small, return of lubrication oil to the compressor 1
deteriorates, and the problem of insufficient lubrication of the
compressor 1 is susceptible to occurrence.
Accordingly, special modifications were priorly made to the control
characteristics and so on of the expansion valve 5, so that refrigerant
was returned to the compressor 1 during low load, thereby eliminating
insufficient lubrication of the compressor 1, but according to the present
embodiment, during low load when low-capacity running of the
variable-capacity compressor 1 is performed, liquid-phase refrigerant can
be caused to be bypassed to the gas-phase refrigerant discharge passage 9
by means of regulating the aperture ratio of the two aperture resistances
11 and 12 as described above, and so the effect of easily being able to
return liquid-phase refrigerant containing lubrication oil to the
compressor 1 can also be demonstrated.
Consequently, an expansion valve 5 of standard type can be utilized in
common without the need to utilize a special expansion valve 5.
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